JPH0861001A - Turbine step structure - Google Patents

Abstract

PURPOSE: To provide a turbine step structure having no resonance of a moving blade due to the slipstream of a stator blade by applying a fluid exciting force to the moving blade passing through the field of flow. CONSTITUTION: In a turbine step structure constructed by assembling the stator blades and the moving blades of an axial flow turbine, the moving blades a on a full periphery are continuously connected together, in a circumferential direction by a linking member, etc., and when a disposing pitch in the circumferential direction indicated by an angle from the rotor center of the moving blade is PB, a disposing pitch PN in the circumferential direction indicated by an angle from the rotor center of the stator blade on the upstream side of the moving blade is set larger than 4PB/3 and smaller than 2PB. Thus, the resonance of the moving blade due to the slipstream of the stator blade is prevented and safety and reliability can be improved.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a turbine stage structure in which a stationary blade and a moving blade of an axial flow turbine such as a steam turbine or a gas turbine are combined, and more particularly, to the entire circumference of a structure in which the moving blade is connected by a connecting member. The present invention relates to a combined structure of a moving blade and a stationary blade.

[0002]

2. Description of the Related Art Regarding such a turbine stage structure,
Particularly, as described in JP-A-58-85304 and JP-A-4-284101, the blades are formed by shifting the pitch between the moving blades and the stationary blades.

[0003]

In the prior art, it was not always possible to suppress the resonance of the moving blade, and it was not possible to achieve a highly reliable paragraph structure.

Therefore, the object of the present invention is to eliminate the above-mentioned drawbacks associated with the turbine stage structure, and to prevent the rotor blade from resonating due to the wake of the stationary blade, or even if it resonates, it is safe and reliable. It is to provide a highly novel paragraph structure.

[0005]

The exciting force acting on the rotor blades of an axial flow turbine such as a steam turbine or a gas turbine is a fluid exciting force due to the nonuniformity of the working fluid, or a mechanical force due to an imbalance of the rotating body itself. Exciting force, electromagnetic exciting force, etc. Of these, the fluid excitation force is the most general one, and is the most important because the blade is highly likely to be damaged when the blade resonates. Since the fluid excitation force acts on the rotor blades passing through the flow field due to the non-uniformity of the flow in the circumferential direction, the excitation frequency is an integral multiple of the rotation speed (hereinafter, rotation speed The integral multiple of is called the excitation order j
Represents).

[0006] Among such fluid exciting forces, there is an exciting phenomenon of the moving blade due to the wake of the stationary blade. The excitation frequency is M _N times the number of revolutions, where M _N is the number of stationary blades on the entire circumference, and M _N differs depending on the paragraph of the turbine. To be taken. The magnitude of the fluid excitation force sharply decreases as the excitation order increases, but a large excitation force exists at a specific order such as the number of vanes. The excitation frequency f (Hz) is M _N times the rotation speed, that is,
When the rotation speed is Ω (rps), it is relatively high because f = _MN Ω, and the rotor blades that cause such an excitation are the natural vibrations of the lower major modes among many rotor blades of the turbine. Relatively short blades whose number is close to the excitation frequency, i.e.
These are blades for high- and medium-pressure turbine stages.

In the vibration design of such a moving blade, the natural frequency of the moving blade is set so that the natural frequency of the main mode of the moving blade does not coincide with the excitation frequency due to the wake of the stationary blade within the turbine operating range. Is designed and adjusted, or the number of stationary blades is changed so that the excitation frequency and the natural frequency of the moving blade do not match. The relatively short blades used in the paragraph of high- and medium-pressure turbines are generally of complicated shape close to iron ingot, and it is difficult to analyze the natural frequency accurately, so that the natural frequency does not match the excitation frequency. Although it was not always easy to do so, the present invention has solved these problems.

In the present invention, when the moving blades of the entire circumference of one turbine stage are seamlessly connected by a connecting member to form a one-ring structure of the entire circumference, and when the number of moving blades of the entire circumference is M _B , The number of stationary vanes M _{N in the} paragraph is set to be larger than M _B / 2 and smaller than 3 M _B / 4.

Further, according to the turbine stage structure of the present invention, the rotor blades on the entire circumference are continuously connected by the connecting member in the circumferential direction, and the rotor blades are arrayed in the circumferential direction represented by the angle from the rotor center. When the pitch is p _B , the arrangement pitch p _N in the circumferential direction, which is represented by the angle from the rotor center of the stationary blades on the upstream side of the rotor blade, is made larger than 4 p _B / 3 and smaller than 2 p _B. It is characterized by

The load operating speed range is outside the range where the resonance of the mode of the node diameter numbers M _B / 4 to M _B / 2 due to the wake of the stationary blade occurs, that is, the rotor operating speed under load operation is Ω (H
z), the natural frequency for the nodal diameter number K (when M _B is the K = M _B / 2, M _B when the even odd _{K = (M B -1) /} 2) and f _m (Hz) At this time, it is preferable that f _m / K <Ω.

[0011]

In the past, the natural frequency of the moving blades of axial turbines such as steam turbines and gas turbines was adjusted by connecting the blades arranged in the circumferential direction to each other except changing the size and shape of the blades themselves. This is often done by using a structure in which they are connected. There are various types of blade connecting structures, and a connecting structure that has been conventionally used is shown in FIG. 2 for an example of an axial flow turbine blade. Further, there is a structural distinction between the case where these connecting members are provided with a circumferential cut and the case where they are not provided. When the connecting member is provided with a cut in the circumferential direction, the blade structure connected by the connecting member is called a finite group blade structure or simply a group blade structure. On the other hand, when the connecting member is not provided with a break in the circumferential direction, the blades on the entire circumference are all connected by the connecting member, and such a blade structure is referred to as a full-circle 1-ring blade structure. As described above, when considering the vibration design of the turbine blade, the natural frequency and the vibration mode also change depending on the blade structure described above.

Here, a brief description will be given of the natural vibration mode of the blade. One blade has a bending or torsional natural vibration mode similar to an ordinary cantilever, but each natural vibration mode has a natural vibration mode group by being connected as a group blade. The number of natural vibration modes of the group blade for each natural vibration mode of one blade is equal to the number of connected blades. For example, in FIG.
The natural vibration modes of five bendings of a five-spell group wing with respect to the natural vibration mode of one bending are shown. On the other hand, the all-round 1-ring blade has a natural vibration mode called a nodal diameter mode, which is similar to the vibration of the disk due to the coupling of the vibrations of the all-round blade. For example, in FIG. The natural vibration modes of the blade diameters 0 to 4 are shown for a blade having 48 blades. Further, when the number of node diameters is represented by k, the number of node diameter modes is 0 ≦ k ≦ K (when M is an even number, K = M / 2, M
Is an odd number, K = (M-1) / 2).

On the other hand, the fluid excitation force acting on such a moving blade and a moving blade structure has a frequency that is an integral multiple of the rotational speed as described above, but in terms of vibration design, the blade is within the operating range of the turbine. So that the natural vibration mode of does not resonate with the excitation force,
Alternatively, it is an important issue to prevent the wings from breaking even if they resonate.

Here, the vibration characteristics of the blade when such a fluid excitation force acts will be examined. In the case of a group wing, if the excitation frequency and the natural frequency match, the magnitude of the vibration response itself will differ depending on the specific configuration of the group wing, the shape of the vibration mode, and the excitation order, but it is basically considered to resonate. There must be. In other words, if the natural frequency of one eigenmode of the group blade and the excitation frequency match, it is necessary to design the vibration by considering that the resonance will occur regardless of the excitation force of any excitation order.

On the other hand, in the case of a full-circle one-ring blade, the resonance condition must satisfy not only the excitation frequency and the natural frequency but also the condition that the eigenmode having the node diameter number k resonates. .

J ± k = λM (1) Here, λ: 0 or a positive integer k: Nodal diameter number (0
≦ k ≦ M / 2) M: number of blades in the entire circumference For example, when λ = 0, j = k, and when λ = 1, j
It shows that resonance occurs when −k = M.

Now, based on the above knowledge, the excitation order j of the excitation force due to the stationary blade wake is j = M _N , and the high-intermediate-pressure turbine having the natural frequency of the low-order main mode close to the excitation frequency. Considering the vibration problem of a relatively short blade for use in a vehicle, when the connecting structure of the blades is a one-ring structure around the entire circumference and the number of moving blades along the entire circumference is M _B , the number of stationary blades in the paragraph is M _N , M _B /
The operation of the configuration that is larger than 2 and smaller than 3M _B / 4 will be described.

FIG. 5 shows an example showing the relationship between the node diameter number k and the natural frequency when a relatively short moving blade for a high-to-intermediate-pressure turbine has a one-ring structure around the entire circumference. The natural frequencies with the axial mode group are shown. The natural frequency has discrete values with respect to the integer k, but is shown by a solid line connecting the natural frequencies in the figure. In general, the change in the natural frequency with the increase in the knot diameter number k is large in the region where the knot diameter number k is small, and is gentle in the region where the knot diameter number k is large. Therefore, if this region is roughly divided into two regions, it can be considered that it is divided into a region in which the node diameter number k is smaller than M _B / 4 and a region in which it is large. From a different point of view, the former is a region where the natural frequency is low in the former region, and the natural frequency is high in the latter region for each natural frequency group of the close combat mode group and the axial mode group. It will also be an area.

Next, it will be examined how the vibration of the rotor blade having such a full-circle one-ring structure appears when the turbine is operating. FIG. 6 is a diagram showing the natural frequency of the nodal diameter mode of the one-circle circumferential blade with respect to the turbine rotational speed, which is a so-called Campbell diagram. In the example of FIG. 6, in order to explain the concept of the present invention as simply as possible, the number of rotor blades on the entire circumference is smaller than that of an actual turbine.
_B = 20, and the natural frequency of the node diameter k = 0 to M _B / 2 is schematically shown by a broken line. Also, for simplicity, FIG.
As described above, the natural frequencies of both the close combat mode group and the axial mode group are not shown, but the natural frequencies of either mode group are schematically shown. It should be noted that a plurality of slanted lines in the figure are lines that represent frequencies that are integral multiples of the number of revolutions (excitation order j times), and those for j = 1 to 20 are shown. Among the intersections of the line representing the frequency that is an integer multiple of the rotational speed in the figure and the line of the natural frequency of the nodal diameter mode indicated by the broken line, ●,
◯ is a point that satisfies the resonance condition expressed by the above equation (1) when the excitation force exists for all of these j = 1 to 20. From this figure, it can be seen that the resonance point of j = k is in a relatively high rotational speed region and the resonance point of j−k = M _B is in a low rotational speed region.

Now, the excitation by the wake of the stationary blade will be considered. In the present invention, (M _B / 2) = 10 <M _N <(3M _B / 4)
= 15, the corresponding range of j = M _N is indicated by hatching in the figure. First, by setting (M _B / 2) <M _N , the excitation order j = M _N due to the wake of the stationary blade becomes larger than the upper limit k = (M _B / 2) of the nodal diameter number of the moving blade. Among the resonance conditions shown in the equation (1), the resonance of j = k can be prevented from occurring.

Next, if (M _B / 2) <M _N , then j + k
It can be seen that a resonance of = M _B may occur. However, since M _N <(3M _B / 4), considering the resonance condition of j + k = M _B , the excited node diameter mode is (M _B / 4) = 5 <k <(M _B / 2) = 10 will be limited. As shown in FIG. 5, the natural frequency of the node diameter mode in this region is 0 <k <(M _B
Although it is in a higher region than the / 4) mode, generally, the higher the frequency is, the larger the kinetic energy required for vibration is, and therefore the vibration response becomes smaller even when the excitation force of the same amplitude acts. There is. The above is a description of the case where a large exciting force due to the stationary blade wake exists in any rotational speed range of FIG. 6, but the exciting force needs to be discussed in relation to the load operating range of the turbine. In other words, when there is no load on the turbine, the flow rate of the working fluid is very small and the fluid exciting force itself due to the wake of the stationary blade becomes small. It is considered that no stress is generated. For example, in FIG. 6, if the rotation speed is the rated rotation speed of the turbine and the rotation speed is within the load operation range, that is, the rotor rotation speed is Ω (H
z), nodal diameter number K (M _B is even when K = M _B / 2, when M _B is an odd number _{K = (M B -1) /} 2 the natural frequency f _m (H) of
z), by setting f _m / K <Ω, the exciting force due to the stationary blade wake becomes very small in the rotation speed range corresponding to the above resonance condition of j + k = M _B , and even if resonance occurs. Very small vibration stress. However, when the load operation range extends to a wide range, resonance corresponding to the condition of j + k = M _B can occur, but by setting j = M _N <(3M _B / 4) as described above, Nodal diameter number M
Only _B / 4 <k <smaller mode of M _B / 2 of the vibration stress is to be excited, the generation of resonance stress that leads to breakage easily avoided.

[0022]

Embodiments of the present invention will be described below with reference to the drawings. FIG. 1 shows a first embodiment of the present invention,
FIG. 3 is a plan view of a turbine stage, which is composed of stationary blades and moving blades of the turbine, as viewed from the radial direction, and is shown expanded over a circumference of 360 °. FIG. 7 is a plan view of a turbine paragraph viewed from the radial direction when the number of moving blades on the entire circumference is 20 as a specific example of the first embodiment of the present invention. Further, FIG. 8 is a perspective view of a portion showing a turbine stage of the present invention.

In FIG. 1, the stationary blades 1 are arranged at equal intervals in the circumferential direction at a pitch p _N expressed as an angle. Similarly,
The rotor blades 2 are also arranged at equal intervals in the circumferential direction at a pitch p _B represented by an angle. Here, the adjacent moving blades 2 are in contact with each other at the end surface 4 of the connecting member 3 and are connected seamlessly over the entire circumference. In the present invention, the stator blade pitch and p _N
The moving blade p _B is set so as to satisfy (4p _B / 3) <p _N <2p _B (2). The reciprocal of the pitch of each of the stationary blades and the moving blades is the total number of the blades of the stationary blades and the moving blades M _N , M _B
Therefore, when the equation (2) is expressed using M _N and M _B ,
It looks like this:

(M _B / 2) <M _N <(3M _B / 4) (3) As a result, (M _B / 2) <M _N as described above in detail regarding the operation of the present invention. By doing so, it is possible to prevent the resonance of j = k due to the excitation order j = M _N due to the vane wake from the resonance conditions shown in the equation (1).
_{By setting B} / 2) <M _N <(3M _B / 4), even if resonance of j + k = M _B may occur, the excited node diameter mode is (M _B / 4) < k <(M _B / 2)
And the natural frequency is relatively high and therefore in the region,
Generally, the higher the frequency is, the larger the kinetic energy required for the vibration is. Therefore, it is possible to limit to the node diameter mode in which the vibration stress is very small even when the resonance occurs.

Further, the rotor rotational speed under load operation of the turbine is Ω (Hz), and the node diameter number K (when M _B is an even number, K = M _B
/ 2, when odd, when the natural frequency for K = (M _B −1) / 2) is f _m (Hz), by setting f _m / K <Ω, the resonance of j + k = M _B can be obtained. It is possible to prevent the occurrence of.

Next, a concrete description will be given with reference to FIG. 7, which shows a case where the number of moving blades on the entire circumference is 20. However, in FIG. 7, the connecting member 3 is not shown. In FIG. 7, the rotor blades 1B to 20B have a pitch p _B represented by an angle in the circumferential direction.
Are evenly spaced. On the other hand, the stationary blades are arranged at equal intervals in the circumferential direction at a pitch p _N represented by an angle of 1N to 13N. That is, p _N = 360 ° / 1
3, p _B = 360 ° / 20, which satisfies the relationship of the above-mentioned formula (2). Nodal diameter number k in nodal diameter mode when 20 rotor blades have a 1-ring structure around the entire circumference
Is in the range of 0 to 10 and the excitation order by the nozzle wake is j = 13, so that resonance that satisfies j = k does not occur. However, for the mode of k = 7, resonance of j + k = M _B = 20 may occur, but in general, the natural frequency of the mode of k = 7 is higher than that of the mode of k = 5 or less. Since it is high, the kinetic energy required for vibration is large, and therefore the vibration stress can often be made small even when resonating. Furthermore, the rotor rotation speed under load operation of the turbine is Ω (Hz), and the number of node diameters is K = M _B /
When the natural frequency for 2 = 10 is f _m (Hz), by setting f _m / K <Ω, the above j + k = M _B
It is also possible to prevent the resonance of.

Next, a connecting member 3 for connecting the moving blades to each other.
8 is a plate-like connecting member that extends integrally with the moving blade from the tip portion of the moving blade 2 in the circumferential direction as shown in FIG. 8, and adjacent moving blades contact each other at the end surface 4 of the connecting member 3. Then, it is like being connected seamlessly over the entire circumference. Here, a brief supplementary explanation will be given on the fact that "they are in contact with each other and are seamlessly connected". First, although the concept of contact has a width, it refers to a state in which the end faces 4 of the connecting members 3 of adjacent moving blades are pressed against each other so that a compressive stress is generated, and the state in which the compressive stress is eliminated is in contact. It is considered that there is no break, that is, there is a break. Therefore, in the example shown in FIG. 7, the end face 4 is in contact with the connecting member 3 by applying some pressing force from the circumferential direction. For example, Fig. 9 shows 2
The state where 0 moving blades 2 are in contact with each other at the end surface 4 of the connecting member 3 and are continuously connected over the entire circumference is shown. However, in order to ensure the contact state, for example, all or some moving blades are shown. It is possible to make the pitch of the connecting members 3 in the circumferential direction slightly larger than a predetermined pitch and to generate a compressive stress in the circumferential direction when the moving blade is assembled to the disk. However, the method of seamlessly connecting the rotor blades of the entire circumference to the one ring is not limited to this, and for example, a method of forming the entire circumference of the one ring structure by using the various connection structures shown in FIG. Of course, it is possible. Of these, a supplementary description will be given below regarding (c) and (f) of FIG. It is well known that, in the case of a blade having a relatively long blade length and being twisted along the radial direction, centrifugal force during rotation causes the blade to be untwisted.
Regarding this, an anzist phenomenon of the tip cover 7 and the intermediate rod 10 and a state in which the blades are connected by bringing the rods into contact with each other and restraining the anzist to connect the blades are shown in a plan view from the outer peripheral side in the radial direction of the blades. Figure 10 and Figure 1
It is 1.

In FIG. 10 (a), the gap 8 between the covers becomes smaller when the untwist occurs, and eventually the gap 8 disappears as shown in FIG. 10 (b), that is, the zero gap 9, and the cover and the cover. Will come into contact and the Anzist will be detained.

The same applies to the case of connection by the intermediate rod 10 shown in FIG. In addition, FIG.
In FIG. 11, the state where there is the gap 8 is a state where there is a disconnection of the connection, and it goes without saying that the state where there is no disconnection of the connection becomes the state of the zero gap 9.

In addition to the above, as means for seamlessly connecting the blades on the entire circumference, as shown in FIG. 12, a plate-like connection extending from the tip of the moving blade 1 integrally with the moving blade in the circumferential direction is provided. The end surface 4 of the member 3 may be provided with a pin hole 11 which extends over both connecting members of adjacent blades and extends along the end surface 4, and the pin 12 may be inserted therein. In this case, both connecting members of the adjacent rotor blades do not necessarily have to be in contact with each other at the end surface 4, and the frictional force generated by the pin 12 being pressed against the pin hole 11 by the centrifugal force acting on the pin 12 during rotation. It may be connected by.

Next, an example in which the turbine stage structure of the present invention is applied to a turbine will be described. FIG. 13 is a sectional view showing a portion of a turbine to which the turbine stage structure of the present invention is applied. An example is shown in which a turbine paragraph including five vanes 1 attached to the inner casing 14 of the turbine and rotor blades 2 attached to the disk 5 provided on the turbine shaft 13 is composed of five paragraphs. The reliability of the turbine is impaired even if any of the turbine paragraphs is damaged by vibration, etc., so that a structure that does not cause a problem even when it is excited by the nozzle wake, which is the main subject of the present invention, is applied to all turbine paragraphs. I need to put it. On the other hand, it goes without saying that the turbine paragraph of the present invention may be applied to all turbine paragraphs, but since not all turbine paragraphs have the above-mentioned reliability problems, the present invention It may be a partial application of the turbine paragraph.

[0032]

As explained above, according to the present invention, there is no resonance of the moving blade due to the wake of the stationary blade, or there is a novel paragraph structure which is safe and highly reliable even if it resonates. It can be provided and can contribute to high reliability of the turbine.

[Brief description of drawings]

FIG. 1 is a developed plan view of a turbine paragraph structure showing an embodiment of the present invention as viewed from a radial direction.

FIG. 2 is a perspective view showing examples of various general blade connection structures.

FIG. 3 is a perspective view showing an example of natural vibration modes of a group blade.

FIG. 4 is a perspective view showing an example of a natural vibration mode of an all-circle one-ring blade.

FIG. 5 is a schematic diagram showing the relationship between the number of node diameters and the natural vibration of a 1-ring blade all around.

FIG. 6 is a schematic diagram showing vibration characteristics during rotation of an all-round single ring blade.

FIG. 7 is a developed plan view of a turbine paragraph structure showing an embodiment of the present invention as viewed from the radial direction.

FIG. 8 is a perspective view of a portion of a turbine stage structure showing an embodiment of the present invention.

FIG. 9 is a front view showing an example of a full-circle one-ring blade structure as seen from the turbine axial direction.

FIG. 10 is a plan view showing an example of a blade connecting structure.

FIG. 11 is a plan view showing an example of a blade connecting structure.

FIG. 12 is a perspective view of a portion showing an example of a blade connecting structure.

FIG. 13 is a sectional view showing a portion of a turbine to which a turbine stage structure of the present invention is applied.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 ... Stationary blade, 2 ... Moving blade, 3 ... Blade connecting member, 4 ... End surface, 5 ...
Disc, 6 ... Blade root, 7 ... Cover, 8 ... Gap, 9 ... Zero gap, 10 ... Rod, 11 ... Pin hole, 12 ... Pin, 13
... shaft, 14 ... inner casing, 15 ... outer casing.

Front page continued (72) Masakazu Takasumi Masakazu Takasumi 502 Jinritsucho, Tsuchiura-shi, Ibaraki Machinery Research Laboratories, Hiritsu Seisakusho Co., Ltd. Hitachi, Ltd.Hitachi factory

Claims (4)

[Claims] 1. A turbine stage structure comprising a combination of a stationary blade and a moving blade of an axial flow turbine, wherein the moving blades of the entire circumference are continuously connected in a circumferential direction by a connecting member, and the entire circumference of the moving blade is connected. A turbine paragraph structure characterized in that, when the number of blades is M _B , the number of blades M _N of the entire circumference of the stationary blade on the upstream side of the blade is larger than M _B / 2. 2. A turbine stage structure comprising a combination of stationary blades and moving blades of an axial flow turbine, wherein the moving blades on the entire circumference are continuously connected by a connecting member in the circumferential direction, and the rotor center of the moving blades is The array pitch in the circumferential direction expressed by the angle
_{When B} is set, the array pitch p _N in the circumferential direction, which is represented by the angle from the rotor center of the stationary blades on the upstream side of the moving blade, is 4p.
Greater than _B / 3, turbine stage structure, which comprises less than 2p _B. 3. A turbine stage structure comprising a combination of a stationary blade and a moving blade of an axial flow turbine, wherein the rotating blades of the entire circumference are continuously connected in a circumferential direction by a connecting member or the like, and the entire circumference of the moving blade is Is M _B , the number of blades M _N of the entire circumference of the stationary blade on the upstream side of the moving blade is larger than M _B / 2, and 3 M _B / 4
Turbine paragraph structure characterized by smaller size. 4. A turbine comprising the turbine stage structure according to claim 1.